EP2189663A2 - Centrifugal compressor - Google Patents
Centrifugal compressor Download PDFInfo
- Publication number
- EP2189663A2 EP2189663A2 EP09176656A EP09176656A EP2189663A2 EP 2189663 A2 EP2189663 A2 EP 2189663A2 EP 09176656 A EP09176656 A EP 09176656A EP 09176656 A EP09176656 A EP 09176656A EP 2189663 A2 EP2189663 A2 EP 2189663A2
- Authority
- EP
- European Patent Office
- Prior art keywords
- blade
- shroud
- camber line
- leading edge
- centrifugal compressor
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Granted
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Classifications
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04D—NON-POSITIVE-DISPLACEMENT PUMPS
- F04D29/00—Details, component parts, or accessories
- F04D29/26—Rotors specially for elastic fluids
- F04D29/28—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
- F04D29/284—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps for compressors
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04D—NON-POSITIVE-DISPLACEMENT PUMPS
- F04D29/00—Details, component parts, or accessories
- F04D29/26—Rotors specially for elastic fluids
- F04D29/28—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
- F04D29/30—Vanes
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- Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
- Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
- Y10T—TECHNICAL SUBJECTS COVERED BY FORMER US CLASSIFICATION
- Y10T29/00—Metal working
- Y10T29/49—Method of mechanical manufacture
- Y10T29/49316—Impeller making
- Y10T29/49329—Centrifugal blower or fan
Definitions
- the present invention relates to a centrifugal compressor provided with a centrifugal impeller, and more particularly to a shape of a blade of the centrifugal impeller.
- centrifugal compressor which compresses a fluid by a rotating impeller (centrifugal impeller) has been widely used for various kinds of plant. Recently, there is a tendency to emphasize a life cycle cost including an operational cost in view of energy (energy saving) and environmental issues, and the centrifugal compressor which has a wide operating range and high efficiency has been expected.
- an operating range of the centrifugal compressor is defined by an area between a surge limit which is a limit on the side of a small flow rate and a choke limit which is an operating limit on the side of a large flow rate.
- a flow rate of gas (working fluid) flowing into the centrifugal compressor is reduced below the surge limit, the centrifugal compressor can not be operated stably by fluctuations of the discharge pressure and flow rate due to separation of flow inside the centrifugal compressor.
- the centrifugal compressor is operated so that the flow rate of the working fluid is between the surge limit and the choke limit.
- JP H10-504621 a technology for improving the efficiency and expanding the operating range by considering a loading distribution of an impeller of a centrifugal compressor is disclosed. Specifically, a generation of a secondary flow inside the impeller is suppressed by concentrating the loading of the shroud side on the leading edge side (upstream side) and the loading of the hub side on the trailing side (downstream side) for expanding the operating range and improving the efficiency.
- the operating range of a centrifugal compressor is further expanded by improving a loading distribution from a leading edge portion (leading edge side of blade) of the shroud side of the impeller to the vicinity of a throat position, and the efficiency (pressure ratio) is further improved, accordingly.
- an object of the present invention to provide a centrifugal compressor provided with an impeller which can improve the efficiency as well as expand the operating range, and further can increase a circumferential velocity.
- a blade angle distribution from a leading edge to a trailing edge of a blade provided in an impeller is determined based on a loading distribution of the blade.
- a centrifugal compressor provided with an impeller which can improve the efficiency as well as expand the operating range, and further can increase a circumferential velocity, can be provided.
- FIG. 1 is a cross sectional view showing a part of a structure of a centrifugal compressor according to a first embodiment of the present invention
- FIG. 2 is a perspective view showing a structure of an impeller.
- a centrifugal compressor 100 includes an impeller 1 which is provided with a blade 7 and rotates around an axis center 5a together with a rotation shaft 5, a diffuser 2 which forms a passage of a working fluid 11, a return bend 3 and a return vane 4.
- the impeller 1, the diffuser 2, the return bend 3 and return vane 4 constitute a single stage and the centrifugal compressor 100 consists of a plurality of the stages arranged in series. That is, a working fluid 11 passed through the return vane 4 in the preceding stage flows into the subsequent stage, and the working fluid 11 is sequentially compressed.
- upstream indicates an upstream of a flow of the working fluid 11 and “downstream” indicates a downstream of the flow of the working fluid 11.
- the impeller 1 is formed in such a manner that a plurality of blades 7 are disposed toward the upstream of a hub 6 which rotates together with the rotation shaft 5 rotating around the axis center 5a.
- a center portion 6a of the hub 6, which is fixed to the rotation shaft 5 gradually expands toward the downstream forming a flange-shape, and the blade 7 which is a plate-like member is vertically disposed along a shape of the hub 6 in the upstream.
- the blade 7 is approximately radially formed toward an edge portion 6b of the hub 6 from a center portion 6a, and a height of the blade 7 is formed to become higher toward the center portion 6a from the edge portion 6b. Meanwhile, the height of the blade 7 is a length from the hub 6 in a direction leaving from the hub 6.
- the blade 7 is formed by such a curved surface that an end of the center portion 6a of the hub 6 is twisted in a rotation direction of the impeller 1.
- a shape of the blade 7 will be described later in detail.
- a shroud 8 which is supported by the blade 7 is provided facing the hub 6, and a plurality of passages 9 surrounded by two blades 7, 7, the hub 6 and the shroud 8 are formed.
- FIG. 2 An illustration where the shroud 8 is partially formed is shown in FIG. 2 . However, this is for showing a shape of the blade 7, and the shroud 8 is provided in entire circumference of the hub 6.
- a side opposite to the hub 6 with respect to the blade in the height direction thereof is called a side of a shroud.
- a flowing velocity of the working fluid 11 flown into the diffuser 2 in FIG. 1 is reduced by a plurality of blades (not shown) and a static pressure is recovered. Then, the working fluid 11 flows into the impeller 1 in the subsequent stage provided in the downstream through the return bend 3 and the return vane 4.
- the flowing velocity of the working fluid 11 is reduced by the plurality of blades, which are not shown, fixed to the diffuser 2, and a loss when the working fluid 11 flows into the return bend 3 can be decreased, thereby resulting in improvement of efficiency of the centrifugal compressor 100.
- the blade 7 includes a camber line (hereinafter, referred to as hub curve line 7b) on a side of the hub 6 and a camber line (hereinafter, referred to as shroud curve line 7a) on the side of the shroud 8.
- hub curve line 7b camber line
- shroud curve line 7a camber line
- leading edge portions a1, b1 End portions of the shroud curve line 7a and the hub curve line 7b in the upstream are named leading edge portions a1, b1, respectively, and those in the downstream are named trailing edge portions a2, b2, respectively.
- An edge connecting the leading edge portion a1 and the leading edge portion b1 forms a leading edge 7L of the blade 7, and the edge connecting the trailing edge portion a2 and the trailing edge portion b2 forms a trailing edge 7T of the blade 7.
- the blade 7 forms a three-dimensional shape where a shape on the side of the hub 6 is defined by the hub curve line 7b and a shape on the side of the shroud 8 is defined by the shroud curve line 7a.
- the shroud curve line 7a and the hub curve line 7b according to the first embodiment are curves which are digitized by the blade angle.
- FIG. 3A is a cross sectional view of an impeller cut at a meridian plane for explaining the blade angle
- FIG. 3B is a cross sectional view of the impeller as seen from the meridian plane
- FIG. 3C is an illustration showing the blade angle.
- a meridian plane Mp at an arbitrary point Pa on the shroud curve line 7a of the blade 7 is a plane including the axis center 5a and passing through the point Pa.
- the meridian plane Mp described above is different depending on a position on the shroud curve line 7a and a position on the hub curve line 7b.
- x shown in FIG. 3A is a length which is measured from the leading edge portion a1 to the point Pa along the shroud curve line 7a, and called as a camber line length.
- a blade angle ⁇ is an angle which is formed between the blade 7 and the meridian plane.
- the blade angle ⁇ between the shroud curve line 7a and the meridian plane and the blade angle ⁇ between the hub curve line 7b and the meridian plane have different values.
- the blade angle ⁇ has a different value depending on a position on the shroud curve line 7a and a position on the hub curve line 7b.
- the blade angle ⁇ (blade angle ⁇ on the side of the shroud curve line 7a) at the point Pa on the shroud curve line 7a of the blade 7 is defined as follows.
- a projected line 7a' is obtained by projecting the shroud curve line 7a on the meridian plane at the point Pa.
- a baseline La on the meridian plane Mp which is tangent to the projected line 7a' at the point Pa is obtained.
- the blade angle ⁇ which is an angle between the baseline La and the blade 7 is formed on a plane orthogonal to the meridian plane Mp at the baseline La.
- a positive direction of the blade angle ⁇ is a rotation direction of the impeller 1 and a negative direction of the blade angle ⁇ is the reverse direction of the rotation direction.
- a distance between the point Pa and the axis center 5a is named as a radius r
- an angle formed between the radius r and a horizontal direction is named as a circumferential direction position ⁇
- a length which is formed by projecting a length between the leading edge portion a1 and the point Pa of the shroud curve line 7a on the meridian plane Mp, that is, a meridional length which is a length of the projected line 7a' shown in FIG. 3B is named as m.
- a shape of the shroud curve line 7a of the blade 7 is determined by continuously setting the blade angle ⁇ (blade angle ⁇ on the side of the shroud curve line 7a) from the leading edge portion a1 to the trailing edge portion a2.
- a shape of the hub curve line 7b is determined by continuously setting the blade angle ⁇ (blade angle ⁇ on the side of the hub curve line 7b) from the leading edge portion b1 to the trailing edge portion b2.
- the blade 7 is formed by smoothly connecting the shroud curve line 7a and the hub curve line 7b, for example, by connecting linearly.
- a shape of the blade 7 formed as described above is an important element which determines a performance of the impeller 1. Therefore, it is required to optimally determine the shape of the blade 7 for obtaining a centrifugal compressor 100 (see FIG. 1 ) which has a wide operating range and high efficiency.
- FIG. 4 is a graph showing a blade loading distribution along a shroud curve line against a non-dimensional camber line length.
- the vertical axis in FIG. 4 indicates a load (blade loading BL) on the blade 7 on the side of the shroud curve line 7a shown in FIG. 2
- the horizontal axis indicates a non-dimensional camber line length S of the shroud curve line 7a shown in FIG. 3C .
- the non-dimensional camber line length S is a non-dimensional number which is calculated by dividing the camber line length x shown in FIG. 3A by a length (whole length) of the shroud curve line 7a.
- the non-dimensional camber line length S is a non-dimensional number which is calculated by dividing a camber line length, which is a length measured along the hub curve line 7b from the leading edge portion b1 to an arbitrary point on the hub curve line 7b, by a length (whole length) of the hub curve line 7b.
- a middle point ct is a point where both the non-dimensional camber lines S of the shroud curve line 7a and the hub curve line 7b become 0.5 (half), and in the shroud curve line 7a, it is a midpoint (midpoint of the shroud curve line 7a) between the leading edge portion a1 and the trailing edge portion a2 along the shroud curve line 7a, and in the hub curve line 7b, it is a midpoint (midpoint of the hub curve line 7b) between the leading edge portion b1 and the trailing edge portion b2 along the hub curve line 7b.
- the blade loading BL is an index indicating a velocity difference and a pressure difference of the working fluid 11 (see FIG. 2 ), which flows on both sides of the blade 7, between both sides of the blade 7, and a velocity reduction rate of the working fluid 11 flowing inside the impeller 1 (see FIG. 2 ) increases as the blade loading BL becomes larger.
- FIG. 5 is a graph showing a relative velocity of a working fluid on a side of a shroud against a non-dimensional camber line length.
- the vertical axis in FIG. 5 indicates a shroud side relative velocity (W/U) calculated as follows.
- An average velocity W is calculated by averaging a relative velocity relative to the blade 7 (see FIG. 2 ) of the working fluid 11 (see FIG. 2 ) on the side of the shroud curve line 7a in the circumferential direction.
- the average velocity W is divided by a circumferential velocity U on the side of the shroud curve line 7a of the impeller 1 (see FIG. 2 ) to calculate the shroud side relative velocity (W/U).
- the horizontal axis indicates a non-dimensional camber line length S of the shroud curve line 7a.
- the shroud side relative velocity (W/U) of the working fluid 11 is a velocity which is obtained by subtracting a circumferential velocity (velocity in circumferential direction) component in the rotation direction of the impeller 1 (see FIG. 1 ) from a main flow velocity of the working fluid 11 in the direction along the rotation shaft 5 (see FIG. 2 ). Since the shroud 8 (see FIG. 2 ) is located on the outer circumferential side and the hub 6 (see FIG. 2 ) is located on the inner circumferential side, a circumferential velocity on the side of the shroud 8 becomes inevitably faster than that on the side of the hub 6.
- the shroud side relative velocity (W/U) on the side of the shroud 8 becomes faster than the relative velocity on the side of the hub 6. Since an aerodynamic loss is substantially proportional to the square of a relative velocity, a relative velocity distribution on the side of the shroud largely effects on a performance of the centrifugal compressor 100 (see FIG. 1 ). Therefore, by optimally designing a shape of the blade 7 on the side of the shroud 8, that is, by optimally designing a shape of the shroud curve line 7a (see FIG. 2 ), a performance of the centrifugal compressor 100 can be secured.
- a blade loading BL along the shroud curve line 7a shown in FIG. 2 linearly goes up at a constant rate from the leading edge portion a1 of the shroud curve line 7a (see FIG. 2 ) as the non-dimensional camber line length S increases, and reaches a maximum value at around the midpoint ct of the non-dimensional camber line length S.
- the blade loading BL decreases linearly at a constant rate as the non-dimensional camber line length S further increases.
- the shroud side relative velocity (W/U) of the working fluid 11 has a maximum value (largest value) at the leading edge portion a1 and then decreases reaching the trailing edge a2 as with the conventional example shown by a dotted line in FIG. 5 .
- the shroud side relative velocity (W/U) of working fluid 11 on the side of the leading edge portion a1 is set larger than that of the conventional example, and the shroud side relative velocity (W/U) at a position distant from the leading edge portion a1 is set smaller than that of the conventional example.
- a distribution of the shroud side relative velocity (W/U) of working fluid 11 was designed such that the shroud side relative velocity (W/U) goes up from the leading edge portion a1 and reaches a maximum value, then, decreases to a value lower than that of the conventional example.
- a throat position is a position at a foot of a perpendicular from the leading edge 7L (see FIG. 2 ) of the blade 7 to the pressure side neighboring blade, in some rotating flow surface (here, shroud surface).
- the shroud side relative velocity (W/U) is large, and if the blade loading BL is large, the shroud side relative velocity (W/U) is small. And, if the blade loading BL along the shroud curve line 7a distributes as shown by the solid line in FIG. 4 , the shroud side relative velocity (W/U) distributes as shown by the solid line in FIG. 5 .
- the blade loading BL on the side of the shroud curve line 7a between the leading edge portion a1 and the vicinity of the throat position is lowered in comparison with the conventional example.
- the leading edge portion a1 is set to a minimum point P MIN of the blade loading BL, and the blade loading BL at the leading edge portion a1 is set to a minimum vale BL MIN .
- a folding point of the distribution of the blade loading BL dominating the blade loading BL from the leading edge portion a1 to the vicinity of the throat position is named P 1 , and the blade loading BL at P 1 is set to BL 1 which can suppress a generation of a reverse flow between the leading edge 7L of the blade 7 and the vicinity of the throat position.
- An optimal value of the BL 1 can be obtained through, for example, experiments.
- the blade loading BL at the leading edge portion a1 and the trailing edge portion a2 may be set to 0 (zero) as long as there is not specific reason.
- the folding point P 1 where a rate of rise of the blade loading BL discontinuously increases is formed between the leading edge portion a1 and the midpoint ct for abruptly increasing the blade loading BL, and the blade loading BL is increased to the maximum value which is larger than that of the conventional example, then, the blade loading BL is decreased toward the trailing edge a2.
- the maximum value in the first embodiment is the maximum value BL MAX of the blade loading BL.
- a point where the blade loading BL has the maximum value BL MAX is named as a maximum point P MAX .
- the folding point P 1 of the blade loading BL may be set, for example, in the vicinity of the throat position of the blade 7 (see FIG. 2 ). That is, it may be possible to distribute the blade loading BL such that the blade loading BL is small at a position between the leading edge portion a1 and the throat position and rapidly increases at a position on the side of the trailing edge portion a2 beyond the throat position.
- setting the blade loading BL 1 at the folding point P 1 to not more than 1/3 of the maximum value BL MAX has the following physical meaning.
- the blade loading BL is 0 (zero) at the leading edge portion a1 and the trailing edge portion a2 and reaches a maximum value at the midpoint ct.
- the throat position is located at around 1/3 from the leading edge portion a1 between the leading edge portion a1 and the midpoint ct in the camber line length x.
- setting the blade loading BL 1 at the folding point P 1 to not more than 1/3 of the maximum value BL MAX means that the blade loading BL is set smaller than the blade loading BL at the throat position in a case when the blade loading BL between the leading edge portion a1 and the midpoint ct is linearly connected. Namely, this indicates that the blade loading BL 1 at the folding point P 1 is set smaller than that of the conventional one.
- setting the blade loading BL 1 at the folding point P 1 to not more than 1/3 of the maximum value BL MAX has the same meaning as securing a surge margin more than ever, and it is preferable to set the blade loading BL 1 at the folding point P 1 to further smaller value for further securing the surge margin.
- a shape of the shroud curve line 7a can be determined using an inverse design method.
- the inverse design method is a method where, for example, a desired distribution of the blade loading BL is calculated first, and subsequently, a shape of the blade 7 is determined based on the distribution. Therefore, the desired distribution of the blade loading BL can be easily realized in comparison with a normal design method, where a shape of the blade 7 is determined first.
- the blade loading BL at the point Pa is a derivative of a product [r ⁇ C ⁇ ], which is a product of the circumferential average absolute velocity C ⁇ and the radius r, differentiated with respect to the camber line length x, and expressed in the next formula (2).
- BL d r ⁇ C ⁇ dx
- the blade loading BL at the point Pa is determined, a relation between the camber line length x and the radius r corresponding to the circumferential average absolute velocity C ⁇ of the working fluid 11 can be calculated. Then, for example, based on the formula (1), the blade angle ⁇ can be set.
- the blade angle ⁇ can be set using the inverse design method, and in addition, by continuously setting the blade angle ⁇ along the shroud curve line 7a, a shape of the shroud curve line 7a can be determined.
- a shape of the hub curve line 7b may be determined using an inverse design method by calculating a desired distribution of the blade loading BL along the hub curve line 7b as with the shroud curve line 7a.
- an effect of the distribution of the blade loading BL along the hub curve line 7b that is, the effect of the distribution of the relative velocity of the working fluid 11 (see FIG. 2 ) along the hub curve line 7b on a performance of the centrifugal compressor 100 (see FIG. 1 ) is smaller than the effect of the distribution of the shroud side relative velocity (W/U) along the shroud curve line 7a.
- a shape of the hub curve line 7b is determined focusing on improvement of strength of the blade 7 shown in FIG. 2 .
- rake angle Le An angle of the trailing edge portion b2 of the hub curve line 7b to be inclined against the trailing edge portion a2 of the shroud curve line 7a is hereinafter called as rake angle Le.
- FIG. 6A is an illustration for explaining a rake angle according to the first embodiment.
- the rake angle Le is an angle between the meridian plane Mp at the trailing edge portion b2 of the hub curve line 7b and the trailing edge 7T.
- the rake angle L ⁇ is an angle between a straight line Lb which is produced by projecting the trailing edge 7T on the meridian plane Mp at the trailing edge portion b2 and the trailing edge 7T, and the rake angle L ⁇ where the trailing edge 7T inclines to a direction to which the impeller 1 rotates is defined as a positive angle.
- the rake angle L ⁇ as defined above is an important index for determining strength of the trailing edge 7T where a stress is the largest in the blade 7. Especially, in the impeller 1 whose circumferential velocity is large or whose pressure ratio is high, the strength of the blade 7 largely depends on the rake angle L ⁇ .
- a shape of the blade 7 is determined by defining the rake angle Le.
- the hub curve line 7b is determined so that an angle between the meridian plane Mp and the leading edge 7L (hereinafter, referred to as leading edge angle Fe) becomes a predetermined angle.
- FIG. 6B is an illustration for explaining a leading edge angle.
- the leading edge angle F ⁇ is an angle between the meridian plane Mp at the leading edge portion b1 and the leading edge 7L.
- the leading edge angle F ⁇ is an angle between a straight line Lc which is produced by projecting the leading edge 7L on the meridian plane at the leading edge portion b1 and the leading edge 7L, and the leading edge angle Fe where the leading edge 7L inclines to a direction to which the impeller 1 rotates is defined as a positive angle.
- the rake angle L ⁇ is set between 0° and +45° and the leading edge angle Fe is set between -10° and +10°, based on the analysis of experiments.
- FIG. 7 is an illustration showing a condition where a weight of a blade is reduced depending on a rake angle.
- the hub curve line 7b is created by connecting the leading edge portion b1 and trailing edge portion b2 so that the blade 7 shown in FIG 2 has a preferable strength and a fluid performance.
- the blade 7 can be created by connecting the shroud curve line 7a and the hub curve line 7b.
- a height of the blade 7 (see FIG. 2 ) can be high. Then, by increasing the height of the blade 7, a passage area of the passage 9 (see FIG. 1 ) can be enlarged, and the centrifugal compressor 100 (see FIG. 1 ) having a large flow rate of the working fluid 11 (see FIG. 2 ) can be configured.
- a flow coefficient suction flow coefficient ⁇ 1 which is an index indicating a flow volume of the working fluid 11 can be set between 0.09 and 0.15.
- the suction flow coefficient ⁇ 1 is a non-dimensional number expressed by the next formula (3), which is inversely proportional to the square of an outer diameter D 2 [m] of the impeller 1 (see FIG. 1 ) and a circumferential velocity U 2 [m/s] of the impeller 1, and proportional to a flow volume (volumetric flow rate) Q [m 3 /s] of the working fluid 11 (see FIG. 1 ).
- ⁇ 1 Q 0.25 ⁇ ⁇ ⁇ D 2 2 ⁇ U 2
- the suction flow coefficient ⁇ 1 expressed by the formula (3) is an index indicating a flow rate of the working fluid 11 flowing in the centrifugal compressor 100 (see FIG. 1 ), and the flow rate of the working fluid 11 can be set larger as the suction flow coefficient ⁇ 1 of the centrifugal compressor 100 becomes larger, thereby resulting in improvement of the efficiency (pressure ratio).
- FIG. 8 is a graph showing a blade angle distribution of a centrifugal compressor according to the first embodiment.
- the vertical axis of FIG. 8 indicates a blade angle ⁇ (The blade angle ⁇ is a negative value according to the definition of the formula (1)) of the blade 7 (see FIG. 2 ), and the horizontal axis indicates the non-dimensional camber line length S.
- FIG. 8 a shape of the blade 7 of the impeller 1 shown in FIG. 2 will be explained.
- a blade angle ⁇ on the side of the shroud curve line 7a is small in the vicinity of the leading edge portion a1, and has a minimum value (minimum value a MIN ) at a position between the leading edge portion a1 and the midpoint ct.
- the blade angle ⁇ on the side of the shroud curve line 7a increases from the minimum value a MIN and has a maximum value (maximum value a MAX ) at a point between the midpoint ct and trailing edge portion a2, then, decreases toward the trailing edge portion a2.
- a change of the blade angle ⁇ in the vicinity of the leading edge portion a1 becomes small, and as shown by the solid line in FIG. 4 , this corresponds to a small blade loading BL in the vicinity of the leading edge portion a1.
- this corresponds to a small change of a flowing direction of the working fluid 11 flowing into the impeller 1 shown in FIG. 1 . Therefore, at the leading edge portion a1, a velocity of the working fluid 11 flown into the impeller 1 may be maintained, or accelerated a little, and accordingly, a surge occurrence at the leading edge portion a1 can be delayed. Namely, a surge limit can be decreased, and an operating range of the centrifugal compressor 100 can be expanded.
- the blade angle ⁇ is rapidly increased at a position from 0.3 to 0.5 of the non-dimensional camber line length S, which corresponds to the vicinity of the throat position.
- the rapid increase of the blade angle ⁇ corresponds to the blade loading BL before and after the folding point P1 shown by the solid line in FIG. 4 .
- An area having a large blade loading BL is an area where a velocity of the working fluid 11 (see FIG. 2 ) rapidly decreases, and the velocity of the working fluid 11 can be decreased in the upstream close to the leading edge portion a1.
- the maximum value (maximum value a MAX ) of the blade angle ⁇ on the side of the shroud curve line 7a, which is located at a position between the midpoint ct and the trailing edge portion a2, contributes to improve the efficiency of the centrifugal compressor 100 by the following reasons.
- the shroud side relative velocity (W/U) which largely effects on the efficiency, is decreased in the upstream of the impeller 1 (see FIG. 1 ) as upper side as possible.
- a position where the shroud side relative velocity (W/U) is decreased and an amount of the decrease of the shroud side relative velocity (W/U) have a close relation to a position where the blade angle ⁇ on the side of the shroud curve line 7a (see FIG. 2 ) rapidly increases and a gradient of the increase.
- the blade angle ⁇ on the side of the shroud curve line 7a is rapidly increased in the first half (upstream side) of the impeller 1.
- the maximum value (maximum value a MAX ) of the blade angle ⁇ becomes larger when the efficiency is prioritized more.
- the maximum value (maximum value a MAX ) of the blade angle ⁇ appears at a position between the midpoint ct and the trailing edge portion a2 on the side of the shroud curve line 7a (see FIG. 2 ).
- the blade angle ⁇ on the side of the shroud curve line 7a has the minimum value a MIN at the leading edge portion a1, but not limited to this position.
- the blade angle ⁇ on the side of the shroud curve line 7a may have the minimum value a MIN at a position between the leading edge portion a1 and the midpoint ct.
- the blade angle ⁇ of each of the shroud curve line 7a and the hub curve line 7b has the same blade angle ⁇ 2 at the trailing edge portions a2, b2.
- the blade angle ⁇ on the side of the shroud curve line 7a at the trailing edge portion a2 and the blade angle ⁇ on the side of the hub curve line 7b at the trailing edge portion b2 are values to be determined based on the specifications of the centrifugal compressor 100 see FIG. 1 ).
- the blade angle ⁇ on the side of the hub curve line 7b has a minimum value b MIN at the leading edge portion b1.
- the blade angle ⁇ increases toward the midpoint ct and reaches a maximum value (maximum vale b MAX ) at a position between the leading edge portion b1 and the midpoint ct, then, decreases toward the trailing edge portion b2.
- the hub curve line 7b is a curve having a single maximum value at a position between the leading edge portion b1 and the midpoint ct.
- the secondary flow loss of the impeller 1 is a loss caused by a velocity difference between the relative velocity on the side of the shroud 8 (see FIG. 2 ) and the relative velocity on the side of the hub 6 (see FIG. 2 ) of the working fluid 11 (see FIG. 1 ).
- a flow toward the shroud 8 from the hub 6 (secondary flow), which is generated so as to absorb the velocity difference, becomes larger as the velocity difference becomes larger. Due to the secondary flow generated as described above, the secondary flow loss is generated.
- the blade angle ⁇ (minimum value b MIN ) at the leading edge portion b1 and the blade angle ⁇ (blade angle ⁇ 2 ) at the trailing edge portion b2 of the hub curve line 7b (see FIG. 2 ) of the impeller 1 are determined based on the specifications (for example, rotation velocity, flow rate and characteristics of working fluid) of the centrifugal compressor 100 (see FIG. 1 ).
- a velocity difference between the velocity on the side of the hub 6 (see FIG. 2 ) and the velocity on the side of the shroud 8 (see FIG. 2 ) depends on a magnitude of the flow coefficient of the centrifugal compressor 100 (see FIG.1 ).
- the impeller 1 (see FIG. 1 ) having a target flow coefficient of the centrifugal compressor 100 according to the first embodiment since the flow difference at the inlet 9a (see FIG. 2 ) is large, it is required that the blade angle ⁇ on the side of the hub curve line 7b (see FIG. 2 ) has a larger maximum value than the blade angle ⁇ 2 at the trailing edge portion b2 for ideally decreasing the flow difference.
- the blade angle ⁇ on the side of the hub curve line 7b has a distribution having the single maximum value b MAX (maximum value) at a position between the leading edge portion b1 and the midpoint ct, as shown in FIG. 8 .
- the impeller 1 having a high reliability and high efficiency (small secondary flow loss) can be configured.
- the shroud curve line 7a intersects with the hub curve line 7b at a position between the midpoint ct and the trailing edge portions a2, b2. That is, a point where the blade angle ⁇ on the side of the shroud curve line 7a and the blade angle ⁇ on the side of the hub curve line 7b have the same value exists at a position between the midpoint ct and the trailing edge portions a2, b2.
- a magnitude relation between the blade angle ⁇ on the side of the shroud curve line 7a (see FIG. 2 ) and the blade angle ⁇ on the side of the hub curve line 7b (see FIG. 2 ) at the leading edge portions a1, b1 (see FIG. 2 ) and the trailing edge portions a2, b2 (see FIG. 2 ) is determined based on the specifications of the centrifugal compressor 100 (see FIG. 1 ).
- the above-described intersection of the blade angle ⁇ occurs when the efficiency is prioritized in the designing.
- a position where the shroud side relative velocity (W/U) is decreased and an amount of the decrease of the shroud side relative velocity (W/U) have a close relation to a position where the blade angle ⁇ on the side of the shroud curve line 7a (see FIG. 2 ) rapidly increases and a gradient of the increase.
- the blade angle ⁇ on the side of the shroud curve line 7a rapidly increases in the first half (upstream side) of the impeller 1. Considering that the blade angle ⁇ at the trailing edge portion a2 is determined by specifications, the maximum value a MAX of the shroud curve line 7a becomes larger when the efficiency is prioritized more.
- FIG. 9 is a graph showing a performance curve of an impeller.
- the impeller 1 according to the first embodiment can obtain a higher pressure ratio than that of the conventional sample shown by a dotted line.
- the impeller 1 can operate with a smaller flow rate of the working fluid 11 (see FIG. 1 ) without causing an occurrence of a surge in comparison with the conventional example. That is, the surge limit can be decreased.
- a choke limit is a maximum flow rate of the working fluid 11 capable of operating the impeller 1. A value of the choke limit is identical to that of the conventional example.
- an operating range of the centrifugal compressor 100 (see FIG. 1 ) provided with the impeller 1 according to the first embodiment can be expanded.
- a strength of the blade 7 can be increased by suitably setting the rake angle L ⁇ (0° to +45°) at the trailing edge 7T of the blade 7 shown in FIG. 6A and the leading edge angle Fe (-10° to +10°) at the leading edge 7L of the blade 7 shown in FIG. 6B .
- the impeller 1 which can rotate at high speed and which can enlarge the circumferential velocity can be configured.
- a distribution of the blade loading BL along the shroud curve line 7a (see FIG. 2 ) according to the first embodiment has the folding point P 1 at the throat position as shown in FIG. 4 .
- FIG. 10 is a graph showing a blade loading distribution having an inflection point.
- the distribution of the blade loading BL may be the one where the blade loading BL smoothly increases as shown in FIG. 10 .
- the distribution of the blade loading BL can be smoothed by forming the inflection point P 2 as shown in FIG. 10
- a distribution of the blade loading BL of the blade 7 (see FIG. 1 ) in the centrifugal compressor 100 depends on a curvature distribution of a blade surface of the blade 7. Therefore, a shape of the blade surface of the blade 7, where the blade loading BL has the inflection point P 2 as shown in FIG. 10 and distributes smoothly, is smooth, and an aerodynamic loss due to, for example, growing of a boundary layer can be decreased.
- a distribution of the blade angle ⁇ on the side of the shroud curve line 7a is determined based on a distribution of the blade loading BL along the shroud curve line 7a.
- a shape of the blade 7 (shape of shroud curve line 7a) having a desired distribution of the blade loading BL can be easily determined by determining a shape of the shroud curve line 7a from the desired distribution of the blade loading BL, by using an inverse design method.
- the impeller 1 (see FIG. 1 ) provided with the blade 7 having a high strength can be obtained.
- the rake angle L ⁇ shown in FIG. 6A is set to a range from 0° to +45° and the leading edge angle F ⁇ shown in FIG. 6B is set to a range from -10° to +10°, a stress to be generated in the blade 7 can be suppressed and strength of the blade 7 can be improved.
- the centrifugal compressor 100 which is provided with the impeller 1 (see FIG. 1 ) capable of improving the pressure ratio as well as expanding the operating range and further capable of increasing the circumferential velocity by using the blade 7 (see FIG. 1 ) according to the first embodiment can be configured.
- FIG. 11 is a graph showing a blade loading distribution along a shroud curve line against a non-dimensional camber line length according to a second embodiment of the present invention.
- FIG. 12 is a graph showing a blade angle distribution corresponding to a blade loading distribution.
- a distribution of the blade loading BL of the blade 7 (see FIG. 2 ) according to the second embodiment on the side of the shroud 8 (see FIG. 8 ) has a maximum value at a position between the midpoint ct and the trailing edge portion a2 of the non-dimensional camber line length S.
- the blade angle ⁇ on the side of the shroud curve line 7a has a maximum vale a MAX at the trailing edge portion a2 as shown in FIG. 12 , corresponding to that the blade loading BL of the shroud 8 distributes so as to have a maximum value at a position between the midpoint ct and the trailing edge portion a2 as shown in FIG. 11 .
- the blade angle ⁇ at the trailing edge portion b2 of the hub curve line 7b has substantially the same value with the maximum value a MAX . Therefore, the blade angle ⁇ on the side of the hub curve line 7b does not intersect with the blade angle ⁇ on the side of the shroud curve line 7a.
- the blade angle ⁇ on the side of the shroud curve line 7a changes more gradually, and a relative velocity of the working fluid 11 (see FIG. 2 ) on the side of the shroud 8 (see FIG. 2 ) decreases more gradually as a peak of the blade loading approaches the trailing edge portion.
- the surge margin can be expanded. Accordingly, it is possible to substantially expand the surge margin by using the impeller 1 (see FIG. 2 ) provided with the blade 7 (see FIG. 2 ) where the blade loading BL distributes as shown in FIG. 11 and the blade angle ⁇ distributes as shown in FIG. 12 .
- the centrifugal compressors according to the embodiments described above can be designed by adjusting a camber line length x having a maximum value of the blade loading in designing a centrifugal compressor where the blade angle on the side of the shroud distributes so that the blade loading has a minimum value at the leading edge, increases from the minimum value along a camber line on the side of the shroud and reaches a maximum value, and decreases from the maximum value along the camber line on the side of the shroud toward the trailing edge, while maintaining a magnitude of the minimum value of the blade loading so that a reverse flow of the working fluid at the leading edge is suppressed.
- the blade angle ⁇ on the side of the shroud curve line 7a distributes so that the blade angle ⁇ has the maximum value a MAX at a position on the shroud curve line 7a closer to the trailing edge portion a2 by moving the position P MAX of the maximum value BL MAX of the blade loading BL closer to the trailing edge, the blade angle ⁇ on the side of the shroud curve line 7a changes more gradually, and thereby, a relative velocity on the side of the shroud 8 (see FIG. 2 ) of the working fluid 11 (see FIG. 2 ) decreases more gradually. As a result, it becomes possible to design a centrifugal compressor which has a wide operating range.
- the efficiency is prioritized in the designing, it is required that a relative velocity on the side of the shroud 8 (the shroud side relative velocity (W/U)), which largely effects on the efficiency, is decreased in the upstream of the impeller 1 (see FIG. 2 ) as upper side as possible.
- a position where the shroud side relative velocity (W/U) is decreased and an amount of the decrease have a close relation to a position where the blade angle ⁇ on the side of the shroud curve line 7a (see FIG. 2 ) rapidly increases and a gradient of the increase.
- the blade angle ⁇ on the side of the shroud curve line 7a distributes so that the blade angle ⁇ has the maximum value a MAX at a position of the shroud curve line 7a (see FIG. 2 ) closer to the leading edge portion a1 by moving the position P MAX of the maximum value BL MAX of the blade loading BL closer to the leading edge, it becomes possible to design a centrifugal compressor which prioritizes the efficiency.
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Abstract
Description
- The present invention relates to a centrifugal compressor provided with a centrifugal impeller, and more particularly to a shape of a blade of the centrifugal impeller.
- A centrifugal compressor which compresses a fluid by a rotating impeller (centrifugal impeller) has been widely used for various kinds of plant. Recently, there is a tendency to emphasize a life cycle cost including an operational cost in view of energy (energy saving) and environmental issues, and the centrifugal compressor which has a wide operating range and high efficiency has been expected.
- When a centrifugal compressor is operated at a constant rotation speed, an operating range of the centrifugal compressor is defined by an area between a surge limit which is a limit on the side of a small flow rate and a choke limit which is an operating limit on the side of a large flow rate. When a flow rate of gas (working fluid) flowing into the centrifugal compressor is reduced below the surge limit, the centrifugal compressor can not be operated stably by fluctuations of the discharge pressure and flow rate due to separation of flow inside the centrifugal compressor.
- In addition, when the flow rate is attempted to increase more than the choke limit, a velocity of the working fluid inside the centrifugal compressor reaches the sonic speed. Then, the flow rate of the working fluid can not be increased more than the choke limit.
- Therefore, the centrifugal compressor is operated so that the flow rate of the working fluid is between the surge limit and the choke limit.
- For example, in
JP H10-504621 - According to the studies of inventors of the present invention, it was found that the operating range of a centrifugal compressor is further expanded by improving a loading distribution from a leading edge portion (leading edge side of blade) of the shroud side of the impeller to the vicinity of a throat position, and the efficiency (pressure ratio) is further improved, accordingly.
- However, there is no description on the loading distribution from the leading edge portion of the shroud side to the vicinity of the throat position in
JP H10-504621 - In addition, since the strength of the impeller is not studied in
JP H10-504621 - It is, therefore, an object of the present invention to provide a centrifugal compressor provided with an impeller which can improve the efficiency as well as expand the operating range, and further can increase a circumferential velocity.
- For solving the foregoing problems, in a centrifugal compressor according to the present invention, a blade angle distribution from a leading edge to a trailing edge of a blade provided in an impeller is determined based on a loading distribution of the blade.
- According to the present invention, a centrifugal compressor provided with an impeller, which can improve the efficiency as well as expand the operating range, and further can increase a circumferential velocity, can be provided.
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FIG. 1 is a cross sectional view showing a part of a structure of a centrifugal compressor according to a first embodiment of the present invention; -
FIG. 2 is a perspective view showing a structure of an impeller; -
FIG. 3A is a cross sectional view of an impeller cut at a meridian plane for explaining a blade angle; -
FIG. 3B is a cross sectional view of the impeller as seen from a meridian plane for explaining the blade angle; -
FIG. 3C is an illustration showing the blade angle for explaining the blade angle; -
FIG. 4 is a graph showing a blade loading distribution along a shroud curve line against a non-dimensional camber line length; -
FIG. 5 is a graph showing a relative velocity of a working fluid on a side of a shroud against a non-dimensional camber line length; -
FIG. 6A is an illustration for explaining a rake angle according to the first embodiment; -
FIG. 6B is an illustration for explaining a leading edge angle of a rake; -
FIG. 7 is an illustration showing a condition where a weight of a blade is reduced depending on a rake angle; -
FIG. 8 is a graph showing a blade angle distribution of a centrifugal compressor according to the first embodiment; -
FIG. 9 is a graph showing a performance curve of an impeller; -
FIG. 10 is a graph showing a blade loading distribution having an inflection point; -
FIG. 11 is a graph showing a blade loading distribution along a shroud curve line against a non-dimensional camber line length according to a second embodiment of the present invention; and -
FIG. 12 is a graph showing a blade angle distribution corresponding to a blade loading distribution. - Hereinafter, a preferred embodiment of the present invention will be explained by referring to drawings as appropriate.
-
FIG. 1 is a cross sectional view showing a part of a structure of a centrifugal compressor according to a first embodiment of the present invention, andFIG. 2 is a perspective view showing a structure of an impeller. - As shown in
FIG. 1 , acentrifugal compressor 100 includes animpeller 1 which is provided with ablade 7 and rotates around anaxis center 5a together with arotation shaft 5, a diffuser 2 which forms a passage of a workingfluid 11, areturn bend 3 and areturn vane 4. - Although not shown in
FIG. 1 , it is noted that theimpeller 1, the diffuser 2, thereturn bend 3 andreturn vane 4 constitute a single stage and thecentrifugal compressor 100 consists of a plurality of the stages arranged in series. That is, a workingfluid 11 passed through thereturn vane 4 in the preceding stage flows into the subsequent stage, and the workingfluid 11 is sequentially compressed. - Hereinafter, "upstream" indicates an upstream of a flow of the working
fluid 11 and "downstream" indicates a downstream of the flow of the workingfluid 11. - As shown in
FIG. 2 , theimpeller 1 is formed in such a manner that a plurality ofblades 7 are disposed toward the upstream of ahub 6 which rotates together with therotation shaft 5 rotating around theaxis center 5a. For example, acenter portion 6a of thehub 6, which is fixed to therotation shaft 5, gradually expands toward the downstream forming a flange-shape, and theblade 7 which is a plate-like member is vertically disposed along a shape of thehub 6 in the upstream. - The
blade 7 is approximately radially formed toward anedge portion 6b of thehub 6 from acenter portion 6a, and a height of theblade 7 is formed to become higher toward thecenter portion 6a from theedge portion 6b. Meanwhile, the height of theblade 7 is a length from thehub 6 in a direction leaving from thehub 6. - In addition, the
blade 7 is formed by such a curved surface that an end of thecenter portion 6a of thehub 6 is twisted in a rotation direction of theimpeller 1. - A shape of the
blade 7 will be described later in detail. - A
shroud 8 which is supported by theblade 7 is provided facing thehub 6, and a plurality ofpassages 9 surrounded by twoblades hub 6 and theshroud 8 are formed. - It is noted that an illustration where the
shroud 8 is partially formed is shown inFIG. 2 . However, this is for showing a shape of theblade 7, and theshroud 8 is provided in entire circumference of thehub 6. - Meanwhile, an "open impeller" may be possible, where the
passage 9 is formed by twoblades hub 6 without using theshroud 8. - It is noted that, even in the "open impeller", a side opposite to the
hub 6 with respect to the blade in the height direction thereof is called a side of a shroud. - When the working
fluid 11 flowing along therotation shaft 5 reaches aninlet 9a, which is opened to the upstream of thepassage 9, theworking fluid 11 flows into thepassage 9 along theblade 7 by a rotation of theimpeller 1. In addition, a pressure of the workingfluid 11 is increased by the rotation of theimpeller 1, and discharged from anoutlet 9b which is opened to the downstream of thepassage 9. After that, theworking fluid 11 flows into the diffuser 2 shown inFIG. 1 . - A flowing velocity of the
working fluid 11 flown into the diffuser 2 inFIG. 1 is reduced by a plurality of blades (not shown) and a static pressure is recovered. Then, theworking fluid 11 flows into theimpeller 1 in the subsequent stage provided in the downstream through thereturn bend 3 and thereturn vane 4. - As described above, the flowing velocity of the
working fluid 11 is reduced by the plurality of blades, which are not shown, fixed to the diffuser 2, and a loss when the workingfluid 11 flows into thereturn bend 3 can be decreased, thereby resulting in improvement of efficiency of thecentrifugal compressor 100. - As shown in
FIG. 2 , theblade 7 includes a camber line (hereinafter, referred to ashub curve line 7b) on a side of thehub 6 and a camber line (hereinafter, referred to asshroud curve line 7a) on the side of theshroud 8. - End portions of the
shroud curve line 7a and thehub curve line 7b in the upstream are named leading edge portions a1, b1, respectively, and those in the downstream are named trailing edge portions a2, b2, respectively. - An edge connecting the leading edge portion a1 and the leading edge portion b1 forms a
leading edge 7L of theblade 7, and the edge connecting the trailing edge portion a2 and the trailing edge portion b2 forms a trailingedge 7T of theblade 7. - As described above, the
blade 7 according to the first embodiment forms a three-dimensional shape where a shape on the side of thehub 6 is defined by thehub curve line 7b and a shape on the side of theshroud 8 is defined by theshroud curve line 7a. - The
shroud curve line 7a and thehub curve line 7b according to the first embodiment are curves which are digitized by the blade angle. -
FIG. 3A is a cross sectional view of an impeller cut at a meridian plane for explaining the blade angle,FIG. 3B is a cross sectional view of the impeller as seen from the meridian plane, andFIG. 3C is an illustration showing the blade angle. - As shown in
FIG. 3A , a meridian plane Mp at an arbitrary point Pa on theshroud curve line 7a of theblade 7 is a plane including theaxis center 5a and passing through the point Pa. - The meridian plane Mp described above is different depending on a position on the
shroud curve line 7a and a position on thehub curve line 7b. - Meanwhile, x shown in
FIG. 3A is a length which is measured from the leading edge portion a1 to the point Pa along theshroud curve line 7a, and called as a camber line length. - A blade angle β is an angle which is formed between the
blade 7 and the meridian plane. The blade angle β between theshroud curve line 7a and the meridian plane and the blade angle β between thehub curve line 7b and the meridian plane have different values. In addition, the blade angle β has a different value depending on a position on theshroud curve line 7a and a position on thehub curve line 7b. - In the first embodiment, the blade angle β (blade angle β on the side of the
shroud curve line 7a) at the point Pa on theshroud curve line 7a of theblade 7 is defined as follows. - As shown in
FIG. 3B , a projectedline 7a' is obtained by projecting theshroud curve line 7a on the meridian plane at the point Pa. In addition, a baseline La on the meridian plane Mp which is tangent to the projectedline 7a' at the point Pa is obtained. - Then, as shown in
FIG. 3C , the blade angle β which is an angle between the baseline La and theblade 7 is formed on a plane orthogonal to the meridian plane Mp at the baseline La. - It is noted that a positive direction of the blade angle β is a rotation direction of the
impeller 1 and a negative direction of the blade angle β is the reverse direction of the rotation direction. - In addition, as shown in
FIG, 3A , a distance between the point Pa and theaxis center 5a is named as a radius r, an angle formed between the radius r and a horizontal direction is named as a circumferential direction position θ, and a length which is formed by projecting a length between the leading edge portion a1 and the point Pa of theshroud curve line 7a on the meridian plane Mp, that is, a meridional length which is a length of the projectedline 7a' shown inFIG. 3B is named as m. Then, the blade angle β can be expressed in the next formula (1) - A shape of the
shroud curve line 7a of theblade 7 is determined by continuously setting the blade angle β (blade angle β on the side of theshroud curve line 7a) from the leading edge portion a1 to the trailing edge portion a2. Similarly, a shape of thehub curve line 7b is determined by continuously setting the blade angle β (blade angle β on the side of thehub curve line 7b) from the leading edge portion b1 to the trailing edge portion b2. - Accordingly, the
blade 7 is formed by smoothly connecting theshroud curve line 7a and thehub curve line 7b, for example, by connecting linearly. - A shape of the
blade 7 formed as described above is an important element which determines a performance of theimpeller 1. Therefore, it is required to optimally determine the shape of theblade 7 for obtaining a centrifugal compressor 100 (seeFIG. 1 ) which has a wide operating range and high efficiency. -
FIG. 4 is a graph showing a blade loading distribution along a shroud curve line against a non-dimensional camber line length. The vertical axis inFIG. 4 indicates a load (blade loading BL) on theblade 7 on the side of theshroud curve line 7a shown inFIG. 2 , and the horizontal axis indicates a non-dimensional camber line length S of theshroud curve line 7a shown inFIG. 3C . - The non-dimensional camber line length S is a non-dimensional number which is calculated by dividing the camber line length x shown in
FIG. 3A by a length (whole length) of theshroud curve line 7a. Similarly, with respect to thehub curve line 7b, the non-dimensional camber line length S is a non-dimensional number which is calculated by dividing a camber line length, which is a length measured along thehub curve line 7b from the leading edge portion b1 to an arbitrary point on thehub curve line 7b, by a length (whole length) of thehub curve line 7b. - A middle point ct is a point where both the non-dimensional camber lines S of the
shroud curve line 7a and thehub curve line 7b become 0.5 (half), and in theshroud curve line 7a, it is a midpoint (midpoint of theshroud curve line 7a) between the leading edge portion a1 and the trailing edge portion a2 along theshroud curve line 7a, and in thehub curve line 7b, it is a midpoint (midpoint of thehub curve line 7b) between the leading edge portion b1 and the trailing edge portion b2 along thehub curve line 7b. - The blade loading BL is an index indicating a velocity difference and a pressure difference of the working fluid 11 (see
FIG. 2 ), which flows on both sides of theblade 7, between both sides of theblade 7, and a velocity reduction rate of the workingfluid 11 flowing inside the impeller 1 (seeFIG. 2 ) increases as the blade loading BL becomes larger. -
FIG. 5 is a graph showing a relative velocity of a working fluid on a side of a shroud against a non-dimensional camber line length. The vertical axis inFIG. 5 indicates a shroud side relative velocity (W/U) calculated as follows. An average velocity W is calculated by averaging a relative velocity relative to the blade 7 (seeFIG. 2 ) of the working fluid 11 (seeFIG. 2 ) on the side of theshroud curve line 7a in the circumferential direction. The average velocity W is divided by a circumferential velocity U on the side of theshroud curve line 7a of the impeller 1 (seeFIG. 2 ) to calculate the shroud side relative velocity (W/U). The horizontal axis indicates a non-dimensional camber line length S of theshroud curve line 7a. - The shroud side relative velocity (W/U) of the working fluid 11 (see
FIG. 2 ) is a velocity which is obtained by subtracting a circumferential velocity (velocity in circumferential direction) component in the rotation direction of the impeller 1 (seeFIG. 1 ) from a main flow velocity of the workingfluid 11 in the direction along the rotation shaft 5 (seeFIG. 2 ). Since the shroud 8 (seeFIG. 2 ) is located on the outer circumferential side and the hub 6 (seeFIG. 2 ) is located on the inner circumferential side, a circumferential velocity on the side of theshroud 8 becomes inevitably faster than that on the side of thehub 6. Accordingly, the shroud side relative velocity (W/U) on the side of theshroud 8 becomes faster than the relative velocity on the side of thehub 6. Since an aerodynamic loss is substantially proportional to the square of a relative velocity, a relative velocity distribution on the side of the shroud largely effects on a performance of the centrifugal compressor 100 (seeFIG. 1 ). Therefore, by optimally designing a shape of theblade 7 on the side of theshroud 8, that is, by optimally designing a shape of theshroud curve line 7a (seeFIG. 2 ), a performance of thecentrifugal compressor 100 can be secured. - Conventionally, as shown by a dotted line in
FIG. 4 , a blade loading BL along theshroud curve line 7a shown inFIG. 2 linearly goes up at a constant rate from the leading edge portion a1 of theshroud curve line 7a (seeFIG. 2 ) as the non-dimensional camber line length S increases, and reaches a maximum value at around the midpoint ct of the non-dimensional camber line length S. In addition, the blade loading BL decreases linearly at a constant rate as the non-dimensional camber line length S further increases. - If the blade loading BL distributes from the leading edge portion a1 toward the trailing edge portion a2 as with the conventional example shown by the dotted line in
FIG. 4 , the shroud side relative velocity (W/U) of the working fluid 11 (seeFIG. 2 ) has a maximum value (largest value) at the leading edge portion a1 and then decreases reaching the trailing edge a2 as with the conventional example shown by a dotted line inFIG. 5 . - However, from recent study results by the inventors of the present invention, it was found that a reverse flow to be generated at the leading edge portion a1 when a flow rate of the working
fluid 11 was decreased causes an occurrence of a surge. Therefore, for delaying the occurrence of the surge, it is preferable to increase the shroud side relative velocity (W/U) of the workingfluid 11 at the leading edge portion a1 to suppress the reverse flow. - On the other hand, for decreasing a fluid loss of the working
fluid 11 flowing in thepassage 9 of theimpeller 1 shown inFIG. 1 , and for improving the efficiency of thecentrifugal compressor 100, it is preferable that a relative velocity on the side of the shroud 8 (seeFIG. 2 ), which is relatively faster than that on the side of the hub 6 (seeFIG. 2 ), is small. - As described above, if the shroud side relative velocity (W/U) of the working
fluid 11 is used as a standard, a suppressing of the surge occurrence conflicts with improving the efficiency of thecentrifugal compressor 100. - Therefore, in the impeller 1 (see
FIG. 2 ) according to the first embodiment, the shroud side relative velocity (W/U) of workingfluid 11 on the side of the leading edge portion a1 is set larger than that of the conventional example, and the shroud side relative velocity (W/U) at a position distant from the leading edge portion a1 is set smaller than that of the conventional example. - For example, as shown by a solid line in
FIG. 5 , a distribution of the shroud side relative velocity (W/U) of workingfluid 11 was designed such that the shroud side relative velocity (W/U) goes up from the leading edge portion a1 and reaches a maximum value, then, decreases to a value lower than that of the conventional example. - Since the
centrifugal compressor 100 is provided with theimpeller 1, where the shroud side relative velocity (W/U) of workingfluid 11 is distributed as described above, the centrifugal compressor 100 (seeFIG. 1 ) can suppress the occurrence of the surge as well as improve the efficiency. Here, a throat position is a position at a foot of a perpendicular from theleading edge 7L (seeFIG. 2 ) of theblade 7 to the pressure side neighboring blade, in some rotating flow surface (here, shroud surface). - In addition, from a correlation between a distribution of the shroud side relative velocity (W/U) of working fluid 11 (see
FIG. 2 ) along theshroud curve line 7a in the impeller 1 (seeFIG. 1 ) and a distribution of the blade loading BL along theshroud curve line 7a of the blade 7 (seeFIG. 2 ), it was found that, for example, if the shroud side relative velocity (W/U) distributes as shown by the solid line inFIG. 5 , the blade loading BL along theshroud curve line 7a of theblade 7 distributes as shown by the solid line inFIG. 4 . In other words, if the blade loading BL along theshroud curve line 7a of theblade 7 is small, the shroud side relative velocity (W/U) is large, and if the blade loading BL is large, the shroud side relative velocity (W/U) is small. And, if the blade loading BL along theshroud curve line 7a distributes as shown by the solid line inFIG. 4 , the shroud side relative velocity (W/U) distributes as shown by the solid line inFIG. 5 . - That is, it is preferable to lower the blade loading BL between the leading edge portion a1 and the vicinity of the throat position for increasing the shroud side relative velocity (W/U) between the leading edge portion a1 (see
FIG. 2 ) and the vicinity of the throat position so as to suppress a reverse flow of the workingfluid 11 between theleading edge 7L (seeFIG. 2 ) of theblade 7 and the vicinity of the throat position - Then, in the first embodiment, as shown in
FIG. 4 , the blade loading BL on the side of theshroud curve line 7a between the leading edge portion a1 and the vicinity of the throat position is lowered in comparison with the conventional example. The leading edge portion a1 is set to a minimum point PMIN of the blade loading BL, and the blade loading BL at the leading edge portion a1 is set to a minimum vale BLMIN. In addition, a folding point of the distribution of the blade loading BL dominating the blade loading BL from the leading edge portion a1 to the vicinity of the throat position is named P1, and the blade loading BL at P1 is set to BL1 which can suppress a generation of a reverse flow between theleading edge 7L of theblade 7 and the vicinity of the throat position. An optimal value of the BL1 can be obtained through, for example, experiments. In addition, the blade loading BL at the leading edge portion a1 and the trailing edge portion a2 may be set to 0 (zero) as long as there is not specific reason. - In addition, the folding point P1 where a rate of rise of the blade loading BL discontinuously increases is formed between the leading edge portion a1 and the midpoint ct for abruptly increasing the blade loading BL, and the blade loading BL is increased to the maximum value which is larger than that of the conventional example, then, the blade loading BL is decreased toward the trailing edge a2.
- It is noted that the maximum value in the first embodiment is the maximum value BLMAX of the blade loading BL. A point where the blade loading BL has the maximum value BLMAX is named as a maximum point PMAX.
- In this case, it was found through experiments that if a blade loading BL1 at the folding point P1 is lowered to not more than 1/3 of the maximum value BLMAX, the efficiency of the impeller 1 (see
FIG. 1 ) can be increased, and thereby, the efficiency of the centrifugal compressor 100 (seeFIG. 1 ) can be improved. - As shown in
FIG. 4 , it may be possible to set the folding point P1 of the blade loading BL, for example, in the vicinity of the throat position of the blade 7 (seeFIG. 2 ). That is, it may be possible to distribute the blade loading BL such that the blade loading BL is small at a position between the leading edge portion a1 and the throat position and rapidly increases at a position on the side of the trailing edge portion a2 beyond the throat position. With the configuration described above, it is possible to obtain such an ideal relative velocity distribution that a velocity reduction of the working fluid 11 (seeFIG. 2 ) at theinlet 9a of theblade 7 in theimpeller 1, which relates to a surge occurrence, is suppressed, and a velocity of the workingfluid 11 is rapidly decreased in the downstream beyond the throat position. - In addition, setting the blade loading BL1 at the folding point P1 to not more than 1/3 of the maximum value BLMAX has the following physical meaning. For example, as an example of a standard blade loading BL, assume that the blade loading BL is 0 (zero) at the leading edge portion a1 and the trailing edge portion a2 and reaches a maximum value at the midpoint ct. Generally, the throat position is located at around 1/3 from the leading edge portion a1 between the leading edge portion a1 and the midpoint ct in the camber line length x. Therefore, setting the blade loading BL1 at the folding point P1 to not more than 1/3 of the maximum value BLMAX means that the blade loading BL is set smaller than the blade loading BL at the throat position in a case when the blade loading BL between the leading edge portion a1 and the midpoint ct is linearly connected. Namely, this indicates that the blade loading BL1 at the folding point P1 is set smaller than that of the conventional one.
- Then, setting the blade loading BL1 at the folding point P1 to not more than 1/3 of the maximum value BLMAX has the same meaning as securing a surge margin more than ever, and it is preferable to set the blade loading BL1 at the folding point P1 to further smaller value for further securing the surge margin.
- If a distribution of the blade loading BL along the
shroud curve line 7a (seeFIG. 2 ) of theblade 7 is determined as described above, a shape of theshroud curve line 7a can be determined using an inverse design method. The inverse design method is a method where, for example, a desired distribution of the blade loading BL is calculated first, and subsequently, a shape of theblade 7 is determined based on the distribution. Therefore, the desired distribution of the blade loading BL can be easily realized in comparison with a normal design method, where a shape of theblade 7 is determined first. - For example, at a point Pa shown in
FIG. 3A , when a radius is r, a circumferential average absolute velocity of the working fluid 11 (seeFIG. 1 ) is Cθ, and a camber line length is x, the blade loading BL at the point Pa is a derivative of a product [r·Cθ], which is a product of the circumferential average absolute velocity Cθ and the radius r, differentiated with respect to the camber line length x, and expressed in the next formula (2). - Therefore, if the blade loading BL at the point Pa is determined, a relation between the camber line length x and the radius r corresponding to the circumferential average absolute velocity Cθ of the working
fluid 11 can be calculated. Then, for example, based on the formula (1), the blade angle β can be set. - Namely, if the blade loading BL is determined, the blade angle β can be set using the inverse design method, and in addition, by continuously setting the blade angle β along the
shroud curve line 7a, a shape of theshroud curve line 7a can be determined. - A shape of the
hub curve line 7b (seeFIG. 2 ) may be determined using an inverse design method by calculating a desired distribution of the blade loading BL along thehub curve line 7b as with theshroud curve line 7a. - However, as described above, an effect of the distribution of the blade loading BL along the
hub curve line 7b, that is, the effect of the distribution of the relative velocity of the working fluid 11 (seeFIG. 2 ) along thehub curve line 7b on a performance of the centrifugal compressor 100 (seeFIG. 1 ) is smaller than the effect of the distribution of the shroud side relative velocity (W/U) along theshroud curve line 7a. - Then, in the first embodiment, a shape of the
hub curve line 7b is determined focusing on improvement of strength of theblade 7 shown inFIG. 2 . - For example, it is known that a strength of the
blade 7 increases if the trailing edge portion b2 of thehub curve line 7b is inclined at a given angle against the trailing edge portion a2 of theshroud curve line 7a. An angle of the trailing edge portion b2 of thehub curve line 7b to be inclined against the trailing edge portion a2 of theshroud curve line 7a is hereinafter called as rake angle Le. -
FIG. 6A is an illustration for explaining a rake angle according to the first embodiment. As shown inFIG. 6A , the rake angle Le is an angle between the meridian plane Mp at the trailing edge portion b2 of thehub curve line 7b and the trailingedge 7T. In more detail, the rake angle Lθ is an angle between a straight line Lb which is produced by projecting the trailingedge 7T on the meridian plane Mp at the trailing edge portion b2 and the trailingedge 7T, and the rake angle Lθ where the trailingedge 7T inclines to a direction to which theimpeller 1 rotates is defined as a positive angle. - The rake angle Lθ as defined above is an important index for determining strength of the trailing
edge 7T where a stress is the largest in theblade 7. Especially, in theimpeller 1 whose circumferential velocity is large or whose pressure ratio is high, the strength of theblade 7 largely depends on the rake angle Lθ. - Accordingly, in the first embodiment, a shape of the
blade 7 is determined by defining the rake angle Le. - In addition, the
hub curve line 7b is determined so that an angle between the meridian plane Mp and theleading edge 7L (hereinafter, referred to as leading edge angle Fe) becomes a predetermined angle. -
FIG. 6B is an illustration for explaining a leading edge angle. As shown inFIG. 6B , the leading edge angle Fθ is an angle between the meridian plane Mp at the leading edge portion b1 and theleading edge 7L. In more detail, the leading edge angle Fθ is an angle between a straight line Lc which is produced by projecting theleading edge 7L on the meridian plane at the leading edge portion b1 and theleading edge 7L, and the leading edge angle Fe where theleading edge 7L inclines to a direction to which theimpeller 1 rotates is defined as a positive angle. - In the first embodiment, the rake angle Lθ is set between 0° and +45° and the leading edge angle Fe is set between -10° and +10°, based on the analysis of experiments.
-
FIG. 7 is an illustration showing a condition where a weight of a blade is reduced depending on a rake angle. - As shown in
FIG. 6B , a radial direction where a centrifugal force works and a direction of theleading edge 7L approach the same direction if the leading edge angle Fθ is decreased close to 0 (zero) on the side of theleading edge 7L where theblade 7 is high, and a bending stress of thehub curve line 7b at the leading edge portion b1, which is generated because the leading edge portion a1 of theshroud curve line 7a is pulled in the radial direction by the centrifugal force, becomes small. - On the other hand, as shown in
FIG. 7 , with respect to the side of the trailingedge 7T, considering that theimpeller 1 including theblade 7 is cut at a predetermined radius of the circumference and the trailingedge 7T of theblade 7 is inclined to the reverse direction of the rotation direction (blade angle β2 is negative), there is a tendency that a weight of theblade 7 to be supported by the trailing edge portion b2 becomes smaller when the rake angle Lθ is a positive value in comparison with a negative value, thereby resulting in reduction of the stress. - That is, as shown in
FIG. 7 , when the rake angle Lθ of theblade 7 is larger than 0° (positive value), a weight of a portion indicated by dots is reduced in comparison with theblade 7 whose rake angle Lθ is 0°, which is indicated by the dotted line. - It was found that a stress by a total force of a centrifugal force operating on the
blade 7 shown inFIG. 2 , a bending force by the workingfluid 11 and a transmitting force inside theblade 7 can be reduced by setting the rake angle Lθ and the leading edge angle Fθ as described above, and accordingly, theimpeller 1 which can endure a large circumferential velocity and high pressure ratio can be manufactured. - Further, the
hub curve line 7b is created by connecting the leading edge portion b1 and trailing edge portion b2 so that theblade 7 shown inFIG 2 has a preferable strength and a fluid performance. - Hence, as described above, the
blade 7 can be created by connecting theshroud curve line 7a and thehub curve line 7b. - In the
blade 7 which has thehub curve line 7b where the strength is considered, a height of the blade 7 (seeFIG. 2 ) can be high. Then, by increasing the height of theblade 7, a passage area of the passage 9 (seeFIG. 1 ) can be enlarged, and the centrifugal compressor 100 (seeFIG. 1 ) having a large flow rate of the working fluid 11 (seeFIG. 2 ) can be configured. For example, a flow coefficient (suction flow coefficient φ1) which is an index indicating a flow volume of the workingfluid 11 can be set between 0.09 and 0.15. - The suction flow coefficient φ1 is a non-dimensional number expressed by the next formula (3), which is inversely proportional to the square of an outer diameter D2 [m] of the impeller 1 (see
FIG. 1 ) and a circumferential velocity U2 [m/s] of theimpeller 1, and proportional to a flow volume (volumetric flow rate) Q [m3/s] of the working fluid 11 (seeFIG. 1 ). - That is, the suction flow coefficient φ1 expressed by the formula (3) is an index indicating a flow rate of the working
fluid 11 flowing in the centrifugal compressor 100 (seeFIG. 1 ), and the flow rate of the workingfluid 11 can be set larger as the suction flow coefficient φ1 of thecentrifugal compressor 100 becomes larger, thereby resulting in improvement of the efficiency (pressure ratio). -
FIG. 8 is a graph showing a blade angle distribution of a centrifugal compressor according to the first embodiment. The vertical axis ofFIG. 8 indicates a blade angle β (The blade angle β is a negative value according to the definition of the formula (1)) of the blade 7 (seeFIG. 2 ), and the horizontal axis indicates the non-dimensional camber line length S. - Referring to
FIG. 8 , a shape of theblade 7 of theimpeller 1 shown inFIG. 2 will be explained. - First, a shape of the
shroud curve line 7a will be explained. - A blade angle β on the side of the
shroud curve line 7a is small in the vicinity of the leading edge portion a1, and has a minimum value (minimum value aMIN) at a position between the leading edge portion a1 and the midpoint ct. - After that, the blade angle β on the side of the
shroud curve line 7a increases from the minimum value aMIN and has a maximum value (maximum value aMAX) at a point between the midpoint ct and trailing edge portion a2, then, decreases toward the trailing edge portion a2. - As described above, since the blade angle β has a minimum value (minimum value aMIN), a change of the blade angle β in the vicinity of the leading edge portion a1 becomes small, and as shown by the solid line in
FIG. 4 , this corresponds to a small blade loading BL in the vicinity of the leading edge portion a1. - Furthermore, this corresponds to a small change of a flowing direction of the working
fluid 11 flowing into theimpeller 1 shown inFIG. 1 . Therefore, at the leading edge portion a1, a velocity of the workingfluid 11 flown into theimpeller 1 may be maintained, or accelerated a little, and accordingly, a surge occurrence at the leading edge portion a1 can be delayed. Namely, a surge limit can be decreased, and an operating range of thecentrifugal compressor 100 can be expanded. - In addition, the blade angle β is rapidly increased at a position from 0.3 to 0.5 of the non-dimensional camber line length S, which corresponds to the vicinity of the throat position.
- The rapid increase of the blade angle β corresponds to the blade loading BL before and after the folding point P1 shown by the solid line in
FIG. 4 . An area having a large blade loading BL is an area where a velocity of the working fluid 11 (seeFIG. 2 ) rapidly decreases, and the velocity of the workingfluid 11 can be decreased in the upstream close to the leading edge portion a1. By decreasing the velocity of the workingfluid 11 as described above, a fluid loss can be decreased, thereby resulting in improvement of efficiency of the centrifugal compressor 100 (seeFIG. 1 ). - In addition, the maximum value (maximum value aMAX) of the blade angle β on the side of the
shroud curve line 7a, which is located at a position between the midpoint ct and the trailing edge portion a2, contributes to improve the efficiency of thecentrifugal compressor 100 by the following reasons. - When the efficiency is prioritized in designing the centrifugal compressor 100 (see
FIG. 1 ), it is required that the shroud side relative velocity (W/U), which largely effects on the efficiency, is decreased in the upstream of the impeller 1 (seeFIG. 1 ) as upper side as possible. A position where the shroud side relative velocity (W/U) is decreased and an amount of the decrease of the shroud side relative velocity (W/U) have a close relation to a position where the blade angle β on the side of theshroud curve line 7a (seeFIG. 2 ) rapidly increases and a gradient of the increase. Therefore, when the efficiency is prioritized in the designing, the blade angle β on the side of theshroud curve line 7a is rapidly increased in the first half (upstream side) of theimpeller 1. Considering that the blade angle β at the trailingedge 7T (seeFIG. 2 ) of theblade 7 is determined by specifications, the maximum value (maximum value aMAX) of the blade angle β becomes larger when the efficiency is prioritized more. As a result, when the efficiency is prioritized in the designing, the maximum value (maximum value aMAX) of the blade angle β appears at a position between the midpoint ct and the trailing edge portion a2 on the side of theshroud curve line 7a (seeFIG. 2 ). - In
FIG. 8 , the blade angle β on the side of theshroud curve line 7a (seeFIG. 2 ) has the minimum value aMIN at the leading edge portion a1, but not limited to this position. The blade angle β on the side of theshroud curve line 7a may have the minimum value aMIN at a position between the leading edge portion a1 and the midpoint ct. In addition, the blade angle β of each of theshroud curve line 7a and thehub curve line 7b (seeFIG. 2 ) has the same blade angle β2 at the trailing edge portions a2, b2. The blade angle β on the side of theshroud curve line 7a at the trailing edge portion a2 and the blade angle β on the side of thehub curve line 7b at the trailing edge portion b2 are values to be determined based on the specifications of thecentrifugal compressor 100 seeFIG. 1 ). A design, where the blade angle β on the side of theshroud curve line 7a at the trailing edge portion a2 and the blade angle β on the side of thehub curve line 7b at the trailing edge portion b2 have the same blade angle β2, is common. - The blade angle β on the side of the
hub curve line 7b (seeFIG. 2 ) has a minimum value bMIN at the leading edge portion b1. The blade angle β increases toward the midpoint ct and reaches a maximum value (maximum vale bMAX) at a position between the leading edge portion b1 and the midpoint ct, then, decreases toward the trailing edge portion b2. As described, thehub curve line 7b is a curve having a single maximum value at a position between the leading edge portion b1 and the midpoint ct. - This, as will be described later, relates to a reduction of a secondary flow loss of the impeller 1 (see
FIG. 1 ). - The secondary flow loss of the
impeller 1 is a loss caused by a velocity difference between the relative velocity on the side of the shroud 8 (seeFIG. 2 ) and the relative velocity on the side of the hub 6 (seeFIG. 2 ) of the working fluid 11 (seeFIG. 1 ). A flow toward theshroud 8 from the hub 6 (secondary flow), which is generated so as to absorb the velocity difference, becomes larger as the velocity difference becomes larger. Due to the secondary flow generated as described above, the secondary flow loss is generated. - Since the hub 6 (see
FIG. 2 ) is located on an inner side rather than the shroud 8 (seeFIG. 2 ) in the radial direction, a relative velocity on the side of thehub 6 becomes small in general in comparison with the relative velocity on the side of theshroud 8. Therefore, a generation of the secondary flow loss can be suppressed by increasing the relative velocity on the side of thehub 6 close to the relative velocity on the side of the shroud 8 (shroud side relative velocity (W/U)) as early as possible. - Considering that a mass flow is preserved from the
inlet 9a (seeFIG. 2 ) to theoutlet 9b (seeFIG. 2 ) of theblade 7 in theimpeller 1, it may be assumed that a meridional velocity Cm at an arbitrary point on the side of thehub 6 is constant regardless of the blade angle β. In addition, considering that the meridional velocity Cm is equal to a projected component of the relative velocity on the meridian plane Mp (seeFIG. 3A ), a relative velocity of a flow flowing along theblade 7 becomes larger as the blade angle β becomes larger. - On the other hand, the blade angle β (minimum value bMIN) at the leading edge portion b1 and the blade angle β (blade angle β2) at the trailing edge portion b2 of the
hub curve line 7b (seeFIG. 2 ) of theimpeller 1 are determined based on the specifications (for example, rotation velocity, flow rate and characteristics of working fluid) of the centrifugal compressor 100 (seeFIG. 1 ). - Therefore, it is effective for suppressing the secondary flow loss in the
impeller 1 to bring a velocity on the side of the hub 6 (seeFIG. 2 ) close to the velocity on the side of theshroud 8 as early as possible, and accordingly, it is required that after the blade angle β on the side of thehub 6 is rapidly increased in the first half (upstream side) of theimpeller 1, the blade angle β is brought close to the blade angle β (blade angle β2) at the trailingedge 7T (seeFIG. 2 ) - A velocity difference between the velocity on the side of the hub 6 (see
FIG. 2 ) and the velocity on the side of the shroud 8 (seeFIG. 2 ) depends on a magnitude of the flow coefficient of the centrifugal compressor 100 (seeFIG.1 ). In the impeller 1 (seeFIG. 1 ) having a target flow coefficient of thecentrifugal compressor 100 according to the first embodiment, since the flow difference at theinlet 9a (seeFIG. 2 ) is large, it is required that the blade angle β on the side of thehub curve line 7b (seeFIG. 2 ) has a larger maximum value than the blade angle β2 at the trailing edge portion b2 for ideally decreasing the flow difference. - Considering the above, the blade angle β on the side of the
hub curve line 7b has a distribution having the single maximum value bMAX (maximum value) at a position between the leading edge portion b1 and the midpoint ct, as shown inFIG. 8 . By distributing the blade angle β on the side of thehub curve line 7b as described above, theimpeller 1 having a high reliability and high efficiency (small secondary flow loss) can be configured. - The
shroud curve line 7a intersects with thehub curve line 7b at a position between the midpoint ct and the trailing edge portions a2, b2. That is, a point where the blade angle β on the side of theshroud curve line 7a and the blade angle β on the side of thehub curve line 7b have the same value exists at a position between the midpoint ct and the trailing edge portions a2, b2. - A magnitude relation between the blade angle β on the side of the
shroud curve line 7a (seeFIG. 2 ) and the blade angle β on the side of thehub curve line 7b (seeFIG. 2 ) at the leading edge portions a1, b1 (seeFIG. 2 ) and the trailing edge portions a2, b2 (seeFIG. 2 ) is determined based on the specifications of the centrifugal compressor 100 (seeFIG. 1 ). The above-described intersection of the blade angle β occurs when the efficiency is prioritized in the designing. - When the efficiency is prioritized in the designing, it is required that a relative velocity (shroud side relative velocity (W/U)) on the side of the shroud 8 (see
FIG. 2 ), which largely effects on the efficiency, is decreased in the upstream of the impeller 1 (seeFIG. 2 ) as upper side as possible. A position where the shroud side relative velocity (W/U) is decreased and an amount of the decrease of the shroud side relative velocity (W/U) have a close relation to a position where the blade angle β on the side of theshroud curve line 7a (seeFIG. 2 ) rapidly increases and a gradient of the increase. Therefore, when the efficiency is prioritized in the designing, the blade angle β on the side of theshroud curve line 7a rapidly increases in the first half (upstream side) of theimpeller 1. Considering that the blade angle β at the trailing edge portion a2 is determined by specifications, the maximum value aMAX of theshroud curve line 7a becomes larger when the efficiency is prioritized more. - In addition, in view of securing a necessary surge margin, a position where the blade angle β on the side of the
shroud curve line 7a (seeFIG. 2 ) rapidly increases can not be moved to the upstream unnecessarily. - Accordingly, when the design is conducted in consideration of securing a minimum necessary surge margin and prioritizing the efficiency, a point where the blade angle β on the side of the
shroud curve line 7a (seeFIG. 2 ) intersects with the blade angle β on the side of thehub curve line 7b (seeFIG. 2 ) appears at a position between the midpoint ct and the trailing edge portions (a2, b2), as shown inFIG. 8 . - A performance of the impeller 1 (see
FIG. 1 ) provided with the blade 7 (seeFIG. 2 ) which has the above-described shapes of theshroud curve line 7a and thehub curve line 7b was measured. -
FIG. 9 is a graph showing a performance curve of an impeller. As shown by a solid line inFIG. 9 , theimpeller 1 according to the first embodiment can obtain a higher pressure ratio than that of the conventional sample shown by a dotted line. In addition, theimpeller 1 can operate with a smaller flow rate of the working fluid 11 (seeFIG. 1 ) without causing an occurrence of a surge in comparison with the conventional example. That is, the surge limit can be decreased. Meanwhile, a choke limit is a maximum flow rate of the workingfluid 11 capable of operating theimpeller 1. A value of the choke limit is identical to that of the conventional example. - Then, an operating range of the centrifugal compressor 100 (see
FIG. 1 ) provided with theimpeller 1 according to the first embodiment can be expanded. In addition, a strength of theblade 7 can be increased by suitably setting the rake angle Lθ (0° to +45°) at the trailingedge 7T of theblade 7 shown inFIG. 6A and the leading edge angle Fe (-10° to +10°) at theleading edge 7L of theblade 7 shown inFIG. 6B . - Accordingly, the
impeller 1 which can rotate at high speed and which can enlarge the circumferential velocity can be configured. - Meanwhile, a distribution of the blade loading BL along the
shroud curve line 7a (seeFIG. 2 ) according to the first embodiment has the folding point P1 at the throat position as shown inFIG. 4 . However, there may be a distribution without the folding point P1. -
FIG. 10 is a graph showing a blade loading distribution having an inflection point. In theblade 7 according to the first embodiment, since a distribution of the blade loading BL along theshroud curve line 7a is sufficient as long as the blade loading BL rapidly increases in the vicinity of the leading edge portion a1, the distribution of the blade loading BL may be the one where the blade loading BL smoothly increases as shown inFIG. 10 . In this case, the distribution of the blade loading BL can be smoothed by forming the inflection point P2 as shown inFIG. 10 - When the inflection point P2 is formed on the distribution of the blade loading BL along the
shroud curve line 7a (seeFIG. 2 ), it was found through experiments that if the blade loading BL2 at the inflection point P2 is smaller than 1/3 of the maximum value BLMAX of the blade loading BL, the efficiency of the impeller 1 (seeFIG. 1 ) can be improved, and a pressure ratio of the centrifugal compressor 100 (seeFIG. 1 ) can be improved. - A distribution of the blade loading BL of the blade 7 (see
FIG. 1 ) in thecentrifugal compressor 100 depends on a curvature distribution of a blade surface of theblade 7. Therefore, a shape of the blade surface of theblade 7, where the blade loading BL has the inflection point P2 as shown inFIG. 10 and distributes smoothly, is smooth, and an aerodynamic loss due to, for example, growing of a boundary layer can be decreased. - As described above, in the blade 7 (see
FIG. 1 ) of thecentrifugal compressor 100 according to the first embodiment, a distribution of the blade angle β on the side of theshroud curve line 7a (seeFIG. 2 ) is determined based on a distribution of the blade loading BL along theshroud curve line 7a. As a result, an operating range of thecentrifugal compressor 100 can be expanded, and the efficiency and the pressure ratio thereof can be increased, thereby resulting in achievement of the excellent effects. - Accordingly, a shape of the blade 7 (shape of
shroud curve line 7a) having a desired distribution of the blade loading BL can be easily determined by determining a shape of theshroud curve line 7a from the desired distribution of the blade loading BL, by using an inverse design method. - In addition, since the blade angle β on the side of the
hub curve line 7b (seeFIG. 2 ) is determined based on a strength of the blade 7 (seeFIG. 1 ), the impeller 1 (seeFIG. 1 ) provided with theblade 7 having a high strength can be obtained. - Especially, if the rake angle Lθ shown in
FIG. 6A is set to a range from 0° to +45° and the leading edge angle Fθ shown inFIG. 6B is set to a range from -10° to +10°, a stress to be generated in theblade 7 can be suppressed and strength of theblade 7 can be improved. - Namely, the centrifugal compressor 100 (see
FIG. 1 ) which is provided with the impeller 1 (seeFIG. 1 ) capable of improving the pressure ratio as well as expanding the operating range and further capable of increasing the circumferential velocity by using the blade 7 (seeFIG. 1 ) according to the first embodiment can be configured. - Next, a second embodiment of the present invention will be explained. Assuming that a centrifugal compressor and components thereof according to the second embodiment are identical to those of the
centrifugal compressor 100 and components thereof shown inFIG. 1 andFIG. 2 , the explanation will be omitted as appropriate. -
FIG. 11 is a graph showing a blade loading distribution along a shroud curve line against a non-dimensional camber line length according to a second embodiment of the present invention.FIG. 12 is a graph showing a blade angle distribution corresponding to a blade loading distribution. As shown inFIG. 11 , a distribution of the blade loading BL of the blade 7 (seeFIG. 2 ) according to the second embodiment on the side of the shroud 8 (seeFIG. 8 ) has a maximum value at a position between the midpoint ct and the trailing edge portion a2 of the non-dimensional camber line length S. - The blade angle β on the side of the
shroud curve line 7a (seeFIG. 2 ) has a maximum vale aMAX at the trailing edge portion a2 as shown inFIG. 12 , corresponding to that the blade loading BL of theshroud 8 distributes so as to have a maximum value at a position between the midpoint ct and the trailing edge portion a2 as shown inFIG. 11 . In addition, the blade angle β at the trailing edge portion b2 of thehub curve line 7b (seeFIG. 2 ) has substantially the same value with the maximum value aMAX. Therefore, the blade angle β on the side of thehub curve line 7b does not intersect with the blade angle β on the side of theshroud curve line 7a. - As described above, by distributing the blade angle β on the side of the
shroud curve line 7a so that the blade angle β reaches the maximum value aMax at the trailing edge portion a2 of theshroud curve line 7a (seeFIG. 2 ), the blade angle β on the side of theshroud curve line 7a changes more gradually, and a relative velocity of the working fluid 11 (seeFIG. 2 ) on the side of the shroud 8 (seeFIG. 2 ) decreases more gradually as a peak of the blade loading approaches the trailing edge portion. - If the relative velocity of the working fluid 11 (see
FIG. 2 ) on the side of the shroud 8 (seeFIG. 2 ) decreases gradually, the efficiency decreases a little, however, the surge margin can be expanded. Accordingly, it is possible to substantially expand the surge margin by using the impeller 1 (seeFIG. 2 ) provided with the blade 7 (seeFIG. 2 ) where the blade loading BL distributes as shown inFIG. 11 and the blade angle β distributes as shown inFIG. 12 . - The centrifugal compressors according to the embodiments described above can be designed by adjusting a camber line length x having a maximum value of the blade loading in designing a centrifugal compressor where the blade angle on the side of the shroud distributes so that the blade loading has a minimum value at the leading edge, increases from the minimum value along a camber line on the side of the shroud and reaches a maximum value, and decreases from the maximum value along the camber line on the side of the shroud toward the trailing edge, while maintaining a magnitude of the minimum value of the blade loading so that a reverse flow of the working fluid at the leading edge is suppressed.
- If the blade angle β on the side of the
shroud curve line 7a (seeFIG. 2 ) distributes so that the blade angle β has the maximum value aMAX at a position on theshroud curve line 7a closer to the trailing edge portion a2 by moving the position PMAX of the maximum value BLMAX of the blade loading BL closer to the trailing edge, the blade angle β on the side of theshroud curve line 7a changes more gradually, and thereby, a relative velocity on the side of the shroud 8 (seeFIG. 2 ) of the working fluid 11 (seeFIG. 2 ) decreases more gradually. As a result, it becomes possible to design a centrifugal compressor which has a wide operating range. - On the other hand, if the efficiency is prioritized in the designing, it is required that a relative velocity on the side of the shroud 8 (the shroud side relative velocity (W/U)), which largely effects on the efficiency, is decreased in the upstream of the impeller 1 (see
FIG. 2 ) as upper side as possible. A position where the shroud side relative velocity (W/U) is decreased and an amount of the decrease have a close relation to a position where the blade angle β on the side of theshroud curve line 7a (seeFIG. 2 ) rapidly increases and a gradient of the increase. Therefore, if the blade angle β on the side of theshroud curve line 7a distributes so that the blade angle β has the maximum value aMAX at a position of theshroud curve line 7a (seeFIG. 2 ) closer to the leading edge portion a1 by moving the position PMAX of the maximum value BLMAX of the blade loading BL closer to the leading edge, it becomes possible to design a centrifugal compressor which prioritizes the efficiency. - Features, components and specific details of the structures of the above-described embodiments may be exchanged or combined to form further embodiments optimized for the respective application. As far as those modifications are readily apparent for an expert skilled in the art they shall be disclosed implicitly by the above description without specifying explicitly every possible combination, for the sake of conciseness of the present description.
Claims (12)
- A centrifugal compressor provided with an impeller which is configured to have a plurality of blades arranged at a predetermined interval in a circumferential direction of a hub rotating together with a rotation shaft,
wherein a blade angle on a shroud side of the blade distributes to have a minimum value at a position between a leading edge of the blade and a midpoint of a camber line on the shroud side, and a maximum value at a position between the midpoint of the camber line on the shroud side and a trailing edge of the blade; and
wherein a blade angle of the blade on a hub side distributes so as to have a maximum value at a position between a leading edge and a midpoint of a camber line on the hub side. - The centrifugal compressor according to claim 1,
wherein the blade angle on the shroud side has a maximum value at the trailing edge. - The centrifugal compressor according to claim 1 or 2,
wherein the blade angle on the hub side is larger than the blade angle on the shroud side at a position between the leading edge and the midpoint of the camber line on the hub side, and smaller than the blade angle on the shroud side at a part of a position between the midpoint and the trailing edge of the camber line on the hub side. - The centrifugal compressor according to at least one of claims 1-3,
wherein if a blade loading at an arbitrary point of the camber line on the shroud side is a derivative of a product of a circumferential average absolute velocity Cθ and a radius r differentiated with respect to a camber line length x as shown by the following formula,where, r is a radius from an axis center of the rotation shaft at an arbitrary point of the camber line on the shroud side, Cθ is a circumferential average absolute velocity of a working fluid flowing in a passage formed in the impeller, and x is a camber line length which is a length measured along the camber line on the shroud side from the leading edge to the arbitrary point of the camber line on the shroud side,the blade angle on the shroud side distributes such that the blade loading has a minimum value at the leading edge, increases from the minimum value along the camber line on the shroud side and reaches a maximum value, and decreases from the maximum value toward the trailing edge along the camber line on the shroud side, while maintaining a magnitude of the minimum value of the blade loading so that a reversed flow of the working fluid at the leading edge is suppressed. - The centrifugal compressor according to claim 4,
wherein the blade loading increases from the minimum value along the camber line on the shroud side and reaches a maximum value at a position between the leading edge and the midpoint. - The centrifugal compressor according to claim 4,
wherein the blade loading increases from the minimum value along the camber line on the shroud side and reaches a maximum value at a position between the midpoint and the trailing edge. - The centrifugal compressor according to at least one of claims 4-6,
wherein a distribution of the blade loading along the camber line on the shroud side has an inflection point at which a rate of rise of the blade loading changes or a folding point where a rate of rise of the blade loading discontinuously increases at a position between a minimum point of the minimum value of the blade loading and a maximum point of the maximum value of the blade loading, the position being between the leading edge and the midpoint of the camber line on the shroud side. - The centrifugal compressor according to claim 7,
wherein the blade loading at the inflection point or the folding point is not more than 1/3 of the maximum value of the blade loading. - The centrifugal compressor according to claim 7 or 8,
wherein the inflection point is a throat position of the blade. - The centrifugal compressor according to at least one of claims 1- 9,
wherein a suction flow coefficient is in a range from 0.09 to 0.15. - A method for manufacturing a centrifugal compressor provided with an impeller which is configured to have a plurality of blades arranged at a predetermined interval in a circumferential direction of a hub rotating together with a rotation shaft, the method comprising steps of:distributing a blade angle on a shroud side of the blade to have a minimum value at a position between a leading edge of the blade and a midpoint of a camber line on the shroud side, and a maximum value at a position between the midpoint of the camber line on the shroud side and a trailing edge of the blade; anddistributing a blade angle of the blade on a hub side so as to have a maximum value at a position between a leading edge and a midpoint of a camber line on the hub side.
- The method for manufacturing a centrifugal compressor according to claim 11, further comprising a step of:determining a distribution of the blade angle on the shroud side from a distribution of the blade loading along the camber line on the shroud side by using an inverse design method.
Applications Claiming Priority (1)
Application Number | Priority Date | Filing Date | Title |
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JP2008298820 | 2008-11-21 |
Publications (3)
Publication Number | Publication Date |
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EP2189663A2 true EP2189663A2 (en) | 2010-05-26 |
EP2189663A3 EP2189663A3 (en) | 2012-07-04 |
EP2189663B1 EP2189663B1 (en) | 2016-04-27 |
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EP09176656.8A Not-in-force EP2189663B1 (en) | 2008-11-21 | 2009-11-20 | Centrifugal compressor and associated manufacturing method |
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US (1) | US8475131B2 (en) |
EP (1) | EP2189663B1 (en) |
JP (1) | JP5333170B2 (en) |
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WO2019006972A1 (en) * | 2017-07-03 | 2019-01-10 | 广东威灵电机制造有限公司 | Impeller, fan and motor |
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Citations (1)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
JPH10504621A (en) | 1994-06-10 | 1998-05-06 | 株式会社 荏原製作所 | Centrifugal or mixed-flow turbomachinery |
Family Cites Families (8)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
JPS60108596A (en) | 1983-11-17 | 1985-06-14 | Ebara Corp | Impeller for centrifugal compressor |
JPH07117357B2 (en) | 1988-07-28 | 1995-12-18 | 三菱電機株式会社 | Control device |
JPH0237297U (en) * | 1988-09-01 | 1990-03-12 | ||
JPH0735091A (en) * | 1993-07-16 | 1995-02-03 | Shinpei Mizuki | Device for preventing swirl in delivery pipe of centrifugal turbo machinery |
JPH08303381A (en) * | 1995-05-10 | 1996-11-19 | Mitsubishi Heavy Ind Ltd | Centrifugal compressor |
JP3673523B2 (en) | 1995-12-07 | 2005-07-20 | 株式会社 荏原製作所 | Turbomachine and manufacturing method thereof |
JP3876195B2 (en) * | 2002-07-05 | 2007-01-31 | 本田技研工業株式会社 | Centrifugal compressor impeller |
US8308420B2 (en) | 2007-08-03 | 2012-11-13 | Hitachi Plant Technologies, Ltd. | Centrifugal compressor, impeller and operating method of the same |
-
2009
- 2009-11-20 JP JP2009265040A patent/JP5333170B2/en active Active
- 2009-11-20 EP EP09176656.8A patent/EP2189663B1/en not_active Not-in-force
- 2009-11-20 US US12/623,089 patent/US8475131B2/en active Active
Patent Citations (1)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
JPH10504621A (en) | 1994-06-10 | 1998-05-06 | 株式会社 荏原製作所 | Centrifugal or mixed-flow turbomachinery |
Cited By (5)
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EP3205883A1 (en) * | 2016-02-09 | 2017-08-16 | Siemens Aktiengesellschaft | Rotor for a centrifugal turbocompressor |
WO2017137207A1 (en) * | 2016-02-09 | 2017-08-17 | Siemens Aktiengesellschaft | Impeller wheel for a centrifugal turbocompressor |
US10865803B2 (en) | 2016-02-09 | 2020-12-15 | Siemens Aktiengesellschaft | Impeller wheel for a centrifugal turbocompressor |
WO2019006972A1 (en) * | 2017-07-03 | 2019-01-10 | 广东威灵电机制造有限公司 | Impeller, fan and motor |
WO2021007195A1 (en) * | 2019-07-10 | 2021-01-14 | Daikin Applied Americas Inc. | Centrifugal compressor for use with low global warming potential (gwp) refrigerant |
Also Published As
Publication number | Publication date |
---|---|
US8475131B2 (en) | 2013-07-02 |
EP2189663A3 (en) | 2012-07-04 |
JP5333170B2 (en) | 2013-11-06 |
EP2189663B1 (en) | 2016-04-27 |
JP2010151126A (en) | 2010-07-08 |
US20100129224A1 (en) | 2010-05-27 |
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